1. Field of the Invention
The invention generally relates to a seal assembly for turbine engines. Specifically, the invention is an intershaft seal capable of providing a seal between inner and outer shafts rotatable about a common axis.
2. Background
Intershaft ring seals are designed to operate between co-rotating or counter-rotating shafts. A ring seal is typically composed of a carbon graphite material and resides within a generally square-shaped or rectangular-shaped channel, often referred to as a gland, formed between end rings or disposed within an inner shaft. When a ring seal is installed within an intershaft system, the seal exerts a small force onto the inner diameter of the outer shaft. During operation, a properly designed ring seal locks onto the outer shaft and spins within the channel or gland. Radial loads along the ring seal are influenced by the pressure between the shafts and centrifugal forces acting on the seal. Axial loads along the ring seal are influenced by the pressure component. Ring seals are designed to have adequate radial loading to prevent the axial load from pushing the seal onto the low pressure end of the channel. FIGS. 1a and 1b describe two intershaft seals from the prior art. It is understood that the left hand side of the seal assembly is the high pressure side and the seal rings are seated on the low pressure side of the end ring or the gland groove.
In FIG. 1a, an intershaft seal assembly 1 is shown including a ring-shaped seal element 2 and spacer ring 6 disposed between a pair of end rings 3, 4. Seal element 2, spacer ring 6, and end rings 3, 4 are further disposed between an outer shaft 10 and an inner shaft 9. End rings 3, 4 have a generally rectangular-shaped cross section and are secured to the inner shaft 9 via a locking ring 5 threaded onto the inner shaft 9 or other means understood in the art. The seal element 2 contacts the outer shaft 10. A gap 54 is disposed between the seal element 2 and the spacer ring 6. The radial height of the gap 54 is sized to avoid contact between the inner diameter of the seal element 2 and the outer diameter of the spacer ring 6 during radial excursions or run-out of the inner shaft 9 and outer shaft 10. In some embodiments, the spacer ring 6 is secured to the end rings 3, 4 via one or more roll pins 8, each disposed within a pin cavity 7 which traverses the spacer ring 6 and end rings 3, 4.
In FIG. 1b, another intershaft seal assembly 1 is shown including a seal element 2 disposed within a ring-shaped gland 43 along an inner shaft 9. The seal element 2 contacts the outer shaft 10 so as to form a seal between the inner diameter of the outer shaft 10 and outer diameter of the seal element 2. The gland 43 could be machined into the inner shaft 9. The axial length of the gland 43 and width of the seal element 2 are tightly controlled to achieve a very tight axial clearance between the seal element 2 and gland 43 so as to minimize leakage of gas from the system.
In both systems described above, the seal element 2 is further forced into the outer shaft 10 by the centrifugal force and pressure loading conditions so as to rotate with the outer shaft 10. In the first system, end rings 3, 4 rotate with the inner shaft 9 and limit axial translation of the seal element 2 along the inner shaft 9. In the second system, the walls of the gland 43 limit axial translation of the seal element 2.
Forces act on the seal element 2 from all sides. On the low pressure side, a pressure drop occurs from high to low pressure from the inner diameter to outer diameter of the seal element 2. The centrifugal loading on the seal element 2 together with radial pressure loads produces a force at the interface between the seal element 2 and outer shaft 10. Axial translation of the seal element 2 is resisted by the friction between the seal element 2 and outer shaft 10. For the seal element 2 to be in equilibrium, the difference in axial forces acting on the seal element 2 must equal the friction force that opposes translation. The relative motion between the seal element 2 and end rings 3, 4 or gland 43 is the sum of the inner and outer shaft speeds in counter-rotating applications and the difference of the inner and outer shaft speeds in co-rotating applications.
The coefficient of friction or pressure ratio greatly influences the performance of an intershaft seal assembly 1. If the coefficient of friction is not properly designed, the pneumatic force acting on the seal element 2 in the axial direction could overcome the opposing frictional force. Accordingly, the coefficient of friction should be kept below the implied coefficient of friction to avoid interactions between the seal element 2 and end rings 3, 4 or walls of the gland 43. Otherwise, the pressure force will push the seal element 2 against the end rings 3, 4 or gland 43 causing the seal element 2 to wear and overheat.
Wear and heating along the seal element 2 are minimized by planar contact surfaces between the seal element 2 and end rings 3, 4 or gland 43; however, wear and heating remains a significant challenge for most intershaft systems. Fluid films are sometimes used between a seal element 2 and end rings 3, 4 or gland 43 to reduce wear via the introduction of hydrodynamic bearing structures such as Rayleigh pads or spiral grooves along the faces of the seal element 2 or end rings 3, 4; however, such films do not have sufficient strength to overcome loading conditions typically encountered within intershaft seal systems. Furthermore, applications including fluid films must properly balance the leakage of gas within and from the seal system to avoid overheating conditions.
A typical seal element 2 will wear during its break-in period as it contacts the end rings 3, 4 or gland 43. This break-in period is completed when the axial clearance between the seal element 2 and end rings 3, 4 or gland 43 is equal to the combined axial run out of the end rings 3, 4 or gland 43. After the break-in period, the wear rate sharply decreases. However, wear remains a substantial challenge when the relative axial translation between inner shaft 9 and outer shaft 10 is greater than the break-in wear clearance, causing the end rings 3, 4 or gland 43 to “bump” the seal element 2 resulting in one revolution of wear for each axial translation.
As is readily apparent from the discussions above, the related arts do not include an intershaft seal capable of minimizing wear and heating during use.
Accordingly, what is required is a seal system capable of avoiding the wear and temperature problems associated with currently available intershaft seal systems.